Variable speed control



Nov. 12, 1940. 'w. L. CARNEGIE 2,221,393

/ VARIABLE SPEED CONTROL I Filed July 1, 1958 6 Sheets-Sheet l moo r F E 8 1% &

m ll (a a Inventor gill/1am [1 (22712 715 2 l q! (Ittornegs Nov. 12, 1940. w. L. CARNEGiE VARIABLE SPEED CONTROL Filed July 1, 1938 6 Sheets-Sheet 2 Zmoentor QhY/zkzzzz ,C 31mg Nov. 12, 1940. w. L. CARNEGIE 3 VARIABLE SPEED CONTROL Filed Jl lly l, 1958 6 Sheets-Sheet 4 Nov. 12, 1940. w. 1.. CARNEGIE 2,221,393

VARIABLE SPEED CONTROL Filed July 1, 1958 6 Sheets-Sheet 5 Zhwentor gh'l/zkzm Aim 712.5415

Gttornegs Nov. 12, 1940. w. 1.. CARNEGIE 2,221,393

VARIABLE SPEED CONTROL Filed July 1, 1938 s Shets-Sheet e V 265 95,2 260 V2 I I I, V. 1 \I 1 2 2&5

Snventor 8g attorneys Patented Nov. 12, 1940 UNITED STATES PATENT OFFICE 2,221,393 VARIABLE SPEED CONTROL Application July 1, 1938, Serial No. 216,889

Claims.

The present invention relates to motor vehicles,

to driving mechanism therefor, and tocontrols for changing speed ratios. It has for its principal object the provision of improved means for initiating the drive of the vehicle through the agency of manual and automatic controls, by compelling gradual pick-up of the torque load of the vehicle through friction elements normally utilized to establish one of the speed ratios of the driving mechanism or variable speed transmission. The invention herewith relates to drive controls, in particular to variable speed gear having fluid pressure servo actuation, and controlled by fluid pressure devices, automatically operated.

It also is an object of the invention to achieve a method for initiating the drive wherein exceptional advantages in smooth starting and acceleration appear, and wherein coordinate response between speed responsive, torque responsive and driver responsive elements of a control system yield new results in the application of engine power to the inertia and torque of a vehicle.

An important object is the provision of means for obtaining the drive-initiating control effect wherein the combined manual-automatic regulation of drive engagement operates cyclically in such a way that there is no hunting, the drive pick-up being completed at one control speed, and released for disengagement at a lower control speed.

i A further object is the provision of fluid pressure servo means for performing the above noted objects, wherein a maximum of smoothness in operation is obtained.

An additional object is the utilization of an automatic control in which a self-metering hydraulic governor furnishes variable fluid pressure for controlling the graduation of drive engagement referred to preceding.

One of the objects of the invention is the utilization of novel servo actuation means for the friction elements of the transmission, having the dual utilities of the initiation of drive, and establishing of one of the driving speed ratios, where--' in new combinations of self-energizing means with the aforesaid means and controls appear, whereby the above objects are in part accomplished.

A further object is the provision of a fluid servo system affording compound multiple speed ratio selections, while yielding interlocking controls among plural transmission units such that for given manual selections of speed ratio, elements which would normally set up drive through cermm of the trains of the driving mechanism are prevented from acting.

An additional object is the provision of interlocking control means affording positive braking of the vehicle if the engine should stall, or be disconnected from the drive through coupling means within the gearing of the driving mechamsm. v

The invention is particularly applicable to change speed gearing systems wherein the shift between speed ratios occurs normally without release of engine torque, but wherein for starting and stopping purposes it is necessary for such release to be available.

Among other objects, the invention is to provide a form of sequence control for forward speed ratios derived through cross-compounding of drive through multiple series units, wherein entirely manual selection is obtainable with controlled graduation of drive between forward ratios and during the speed-change intervals, without interruption of torque.

A further object is the provision of a pressure compensation action worked by a driver control to establish the net torque-capacity of friction elements in the transmission supporting the drive, coordinated to operate for any of the selected shifts of ratio, on a continuously effective basis. This enables the ratio selector to be moved from any one position to any other, assuring a proper graduation of transfer of drive at all times to avoid shock to passengers or damage to the driving parts.

The invention contemplates a system of drive in which ratio determining elements of variable speed gearing may also be used for initiating and establishing drive automatically, so that the customary main clutch may be dispensed with, and economies in construction and cost derived.

Other objects and advantages will be understood from the following description.

In the accompanying drawings:

Figure 1 is a vertical, longitudinal section of the two speed gear unit of an example of the application of my invention, showing the drive for the servo pump and governor device.

Figure 2 is a transverse, vertical section at 2--2 of Figure 1 showing the energization and control method for the gearing of Figure 1.

Figure 3 is a transverse sectional view at 33 of Figure 1 showing the servo pump and hydraulic governor utilized to control the initiation of drive and-the ratio transitions.

Figure 4 is a schematic diagram illustrating the control valving for the initiation of drive in conjunction with the pump and governor of Figure 3, and the servo and energization system of Figure 2.

Figure 5 is a diagram of a typical governor speed-pressure curve for a governor of the type of Figure 4.

Figure 6 is a corresponding diagram of pressures derived from Figure 5 by the structure of the valving, springs, plungers and pistons of Figures 2 and 4, showing ther'non-hunting characteristic of the arrangement.

Figure 7 is a partial view, in transverse section, of Figure 1 at 1--1, wherein the actuation means for the second unit of the drive assembly is described.

Figure 8 is a schematic diagram of the controls for the gearing of Figure 1 including in outline the control groups of Figures 2, 3, 4 and 7 in conjunction with the actuation means for the reverse drive unit and valves controlling all of the units.

Figure 9 shows a control linkage for the valving of Figure 8, the schematic projection of Figure 10 illustrating the external controls for the elements of Figure 9 and the graduating control operated by the compensator and accelerator-connected linkage of Figure 8.

Figure 11 shows a supplementary control for the band of thereverse unit wherein the operator may apply the band and establish a locking couple within the gearing when engine braking is not available.

The gearv assembly used to illustrate my invention shown in Figure 1 has engine-connected shaft 3 to which is afilxed annulus gear 4 with ring of internal teeth 5 meshing with planet gears 6 spindled on carrier 1. Bearings 8 support sleeve 9 and its sun gear 2| on output shaft IU of the first unit piloted in bearing II in shaft 3.

The carrier 1 drum l2 splined date clutch plates |6.' Clutch plates 20 mating with plates 5 are mounted on drum extension 8 of sleeve 9 and reaction sun gear 2| meshing externally with planets 6. The gear 2| is joined to sleeve 9 and to drum l8, the bearings 8 supporting the assembly on the shaft I0.

Brake band 30 may prevent the rotation of drum l8 when locked. Clutch release springs 22 normally bias plates |52|I for disengagement, by unloading presser plate 23 in tension. The presser plate 23 may also slide longitudinally with plates 20 on key bolts not shown.

The drum I8 is formed into cylindrical recesses 3| for pistons 32, the cylinders 3| communicating through passages 33 with passage 34 in gland 35 fixed against rotation with respect to shaft I0, and having external pressure pipe 50.

Figure 2 shows a section of Figure 1, giving details of external servo controls for the operation of brake band 30.

The band 30 encircles drum l8, and its two ends 36 and 31 are equipped with fixtures 38 and 39. Fixture 38 has notch 4| in which thrust rod 42 is seated. Rod 42 is pivoted at 43 is extended to form a clutch externally at |3 to accommo- 0 to lever 44, pivoted to the casing 12 at 48. Flat 41 at the opposite end of lever 44 accommodates rod 48.

Fixture 39 similarly has notch 5| on which thrust rod 52 is seated. Rod 52 is pivoted at 53 to lever 54, pivoted to the casing 60 at 56. Flat l5 having mating spline teeth 51 at the opposite end of lever 54 accommodates rod 58.

Upon consideration of this structure, one will note that the fixtures 38 and 39 may act interchangeably as anchor points depending on the direction of rotation of drum l8, and the degree of wrapping permitted in brake 30 which may have only one or more than one turn about drum l8 between ends 36 and 31.

The normal hand of rotation of engine connected shaft 3 is shown by arrow :0 of Figure 2, and the ldling'direction of rotation of drum I8 when clutch |5-20 is disengaged, and brake 30 unapplied, is shown by arrow Y.

If brake be locked, the drive will be transcylinder 6| may engage piston 63 only when a given motion of translation occurs on the part of the piston 63. Small piston 65 is recessed in sub-cylinder 66, connected to pipe I00. Fluid pressure is admitted to act on piston 63 through pipe lfll. Rod 58 is moved by piston 63 through stops and 65'.

Rod 58 passes through wall 68a of cylinder 60 to chamber 1|, where it is aflixed to piston 61, between which and the may be admitted through pipe I02.

The left end of cylinder 60 is closed by flanged plate 12 and by plate 13 having recess 14, the extension of which acts as an abutment for rod 58. Cylinder 18 in flanged plate 12 houses piston 11 of rod 48, and is connected to pipe I03.

Springs 18 held by .plate 13 are normally loaded to hold piston 61 to the right, and thereby load brake end 31 through rod 58, lever 54 and thrust rod 52.

Small spring 19 exerts longitudinal pressure between piston 63 and abutment 65' which transfers force to rod 58 through abutment 15.

The parts to the left of Figure 1 show the arrangement of drive between the engine connected shaft 3,,the driving shaft I 0 of the two-speed unit just described, and the pump and governor assemblies.

The engine connected shaft 3 in Figure 3 carries .the master drive gear 18 meshing with gear 8| of vertical shaft 82 and gear 83 of transverse shaft 84. The pump assembly P is located in the lower part of the transmission casing 2 and draws oil from the sump through suction pipe H8, and delivers it -to pressure space I, connected to line )5 and 5, as shown also in Figure 8. The supporting casing 2 is drilled at 2 for pipe ||3-and feeds pump pressure to hollow bore 4 of shaft 84. The fluid pressure metering valve 80 ofgovernor G shown in detail in Figure 4 controls the pressure to outlet pipe I04 according to variations in speed of shaft 84 as determined by the speed of engine shaft Figure 4 shows the sectional view of the parts of the governor assembly.

Governor body 90 is keyed to rotate with transwall 68a fluid pressure verse shaft 94 of Figure 1, and is bored out at 9I and at 92 for valve 80. The port 93 adjacent the centerline of rotation is fed by pump pressure through shaft passage I I4.

Connected ports 94 and 95 are the outlet pressure ports of the governor; port 96 exhausts back to sump 20I, and bore 92 is vented at 91. The portions 80a and 80b of valve 80 are machined to fit bore 9|. and portion 800 to fit bore 92. The undercut spaces between the bosses provide the lap means between the ports for the various conditions of operation. Weight W is integral with valve 80.

At zero or low speed, the position shown, the pump pressure in port 93 is sealed against action. At some given increased speed point, boss 80b of valve 80 uncovers port 93 and pump pressure may pass through the adjacent undercut to port 95 and to port 94. Since the area of boss 800 is greater than that of boss 80b, the tendency for the pressure is to close off the port 93 and oppose the centrifugal action of weight W. As

speed increases, the pressure characteristic is overcome by the speed effect so that the pressure in output line rises, according to a scale such as given in the example of Figure 5.

This varying output pressure is utilized to control the operation of band 30 acting as a starting clutch for the vehicle, as will be understood from the description of the operation to follow.

The valve control grouping at the left of Figure 4 is connected to the hydraulic governor just described, and to the servo control system Figures land 2.

Conveniently located, the valve casing H9 is bored to accommodate a number of plungers ported to coordinate variations in pressure with the various control functions desired.

Mai-n control valve I20 is the selector valve for the action of band 30, and consists of bosses I20a, I20b, I200 and I20d of common diameter to fit bore I2I. At its upper end, port I23 is open to output governor pressure through line I04, the adjacent port I24 is connected through I38 to line I03 joining to cylinder 16 of Figure 2; port I26 is .joined to pipe I05 carrying pump pressure, and likewise to counterport I21. The next port in order is I28 connected to pipe I02 and to cylinder H of Figure 2. The remaining two ports I29 and I3I connect to exhaust, or sump, and also through pipe I25 to the uppermost port I32, open to plunger I33 vented at I34.

In order from the top of thedrawings, plunger I33 in bore I2I may press on sliding pin I36 which may transmit force to plunger I31 in bore I2I', ported at I39 to receive pressure from line I38. Plunger I31, vented at I4I, may press pin I42 against boss Ia of valve I20.

The lower portion of bore I2I is ported to exhaust at I43 and contains spring I44 pressing upward against valve I20 and spring I46 of sliding abutment I41 on rod I48, the collar I41 being arranged .to strike an extension of valve I20 at a given limit of travel.

3 In Figure 3 a pressure regulating valve II6 of known type is shown connecting the pump outlet port II1 with pressure outlet line II5, the metering spring II8 regulating the lift of the valve; the line II9' being connected to the transmission 3 lubrication system.

With the vehicle at standstill, engine idling at about 300 r. p. m., the load shaft I0 is idle, and if a driving couple has been established between the vehicle load and shaft I0, the unit input gear 5 4 will be rotating at engine speed.

Drum I8, because of the reaction of stopped carrier 1, spins backward, in a direction opposite to that of gear 4. The torque reaction force is therefore negative.

Pump pressure is effective through pipe I02 on piston 61 of Figure 2, holding off springs 18, so that band 30 is free from the drum I8.

An increase in engine speed initiates governor pressure to act on valve I20 through pipe I04 and port I23, which compresses spring I44. At about 500 r. p. m. of the engine, valve I20 has been moved toward the bottom of the drawing of Figure 4 far enough to close off port I26, and open port I28 to exhaust port I29, to drain cylinder 1 I.

The valve I20 is just on the point of metering pump line pressure from port I26 to the anchor piston port I24.

During this interval, the cylinder H has been drained through pipe I02, and the springs 18 have forced piston 61 to the end of its stroke, carrying rod 58 and lever 54 with it.

This establishes a very low minimum of slack remaining in band 30.

Further increase in engine speed causes the governor pressure to rise, and valve I20 meters a higher and increasing pressure to anchor piston 11, applying end 36 to band 30, establishing torque reaction in the non-energising direction,

since the spinning direction of the drum is negative, as discussed preceding.

At a given engine speed. in the example, about 1050 r. p. m., the governor is delivering 70 pounds to port I23 of the valve I20, which moves further to compress spring I44 and also spring connecting ports I3I and I21.

This puts full pump pressure on plunger I33 through pipe I25 and the force transmitted through pin I36 to plunger I31 and through pin I42, shifts valve I20 to the limit of travel in the direction to fully compress spring I44 and spring I46 as far as the end of rod I48 will allow.

In the consideration of the operation of the control valve assembly of Figure 4, the areas of plungers I33 and I31 are so taken that control valve I20 is held against the action of springs I44 and I46 whenever the governor G is supplying 16 pounds to port I23. At below this pressure, the springs I 44 and I46 are active and valve I29 starts to move toward the top of the drawing.

Continued diminishing of governor pressure with engine speed causes the valve I20 to cut off port I29 which has been supplying full pump pressure through I25 to plunger I33, and port I32 is open to exhaust through port I29.

The spring I44 then shifts the valve I20 to the upper limit of travel, and ports I26 and I28 are connected, supplying pump pressure to piston 61 releasing band 30 completely. Plunger I33 makes it possible to maintain the loop action of the operating curve of Figure 6, to achieve a fixed hysteresis loop effect.

The diagram of Figure 6 is a typical response curve for the valve action of Figure 4. with respect to the actuation system for band 30 as shown in Figure 2.

It will be noted that at 500 r. p. m. the governor pressure in line I04 and acting on plunger I20 becomes effective to overcome spring I44 and direct the valving to drain cylinder H.

The pressure rises from zero to about 70 pounds at 1,000 r. p. m., giving a graduated loading to end 36 of band 30, to maximum at that latter point.

If the engine be decelerated because of the requirement for traffic maneuvering on the part of the operator, the loading achieved at 1000 r. p. m. is maintained, and the load shaft I0 remains coupled to the engine until the engine speed diminishes to 400 r. p. m., when the governor supplied pressure falls off such that springs I44 and I 46 are fully effective to shift valve I20 to open position, relieving the band 30 from any energising whatever.

This attains a positive neutral or no-drive condition at all times unless the engine stall, wherein pump P will not be able to supply lines I05; I02 and cylinder 60.

In the arrangement of gearing thus far described, the system of starting clutch control involving the functioning of valve I20 is applied to forward drive only, the added arrangement of Figure 3 being necessary to attain a degree of controlled graduated pick-up of drive in reverse, or some equivalent thereof.

Referring back to Figures 2 and 4, the duplication of the parts of cylinder 60 and rods 48, 58, and the control valve assembly of Figure 4 including the action of valve I20 makes it possible to install a simple directing valve in pipe I04 operated by the operators hand lever when shifted to reverse, so that the duplicate set of parts becomes active to control the action of brake end 31, instead of brake end 36 as previously described. This would necessitate two cylinders similar to cylinder 60, one of which would be connected as in Figures 2 and 4, the other of which would have rod 58 acting against lever 44, and rod 48 acting against lever 54.

The necessity for such a duplication of parts is created by the fact that with a reversing .gear placed ahead of the planetary gearbox, the reaction member I8 would idle oppositely to the direction of rotation of the input shaft 3 so that the graduation controls for end 36 of brake 30 worked by valve I20 could not serve for energisation control of end 31 of brake 30.

In order to make the arrangements of Figures 2 and 4 practically useful, I therefore prefer to connect shaft 3 directly to the engine shaft at all times, and utilize a gear structure wherein the reversing gear train is between the described unit and the final output shaft.

It will be seen that the illustrative arrangement of Figure l complies with the above requirement, so that for either forward or reverse drive the reaction element I8 of the unit in which I utilize the reaction brake for a vehicle starting means will always have a negative idling direction of rotation such that brake end 36 may be controlled for graduating initial drive from rest in either forward or reverse, avoiding the self-energising effect which would be harsh and abrupt in function, and not desirable from either a commercial, manufacturing, or a durability point of view.

In the operation of the vehicle with the invention as far as described the operator may start his engine, and warm it up in the regular way without interference by any part of the described mechanism.

When ready to drive forwardly, the operator shifts the manual controls, and it is assumed that shaft I0 is made subject to the load of the vehicle thereby.

The accelerator pedal is depressed, and the vehicle moves forward as the engine power is developed, according to the torque reaction sustaining ability of band 30, which instead of selfenergising, as in common mechanisms of this character, is fluid pressure energised in accordance with the driver's accelerator pedal action, as transferred into terms of engine speed by governor G.

If the process of accelerating is interrupted by trafllc conditions, the cyclic action of initiating drive is repeated between predetermined engine speeds, so that unnecessary slippage or abuse is safeguarded against.

Various forms of reverse gearing may be utilized, but for ease of control and other considerations, the form herewith to be described is believed to offer certain advantages in compactness and ease of control, as well as in the disposition of the servo and automatic control equipment.

Figure 1 represents a typical gear arrangement for overall drive from the engine to the tailshaft of the vehicle, the gearing being shown connected to a second form of planetary unit which in turn is connected to a third gear, the latter being a reversing planetary.

In this figure the shaft I0 is joined to the carrier 1 of the first unit, and to the two pinions I5I and I52 of the second unit, likewise having clutch drum I50 keyed to it. It will be noted that shaft I0 is the output member of the first unit and the power input member of the second unit.

Carrier I56 and shaft 300 are joined together, the carrier being the output member of the second unit. Carrier I12 of the reversing or third unit is likewise attached to or integral with output shaft 300.

As will be described binations for these units are obtained through manipulation of clutch I5-20 and brake 30 of the first unit; clutch I55-I60 and brake I10 of the second unit, and brake I of the third unit.

In Figure 1 shaft 3 is directly connected to the engine, with input gear 4, reaction sun gear 2|, planets 6, carrier 1, reaction gear sleeve 9 and drum I8. The output shaft Io extended to the right carries the two pinions or sun gears I5I and I52, and clutch drum I50, with a splined set of clutch plates I55. Pinions I5II52 are the input gears of the second unit.

The pinion I 52 meshes with a set of planets I 54 spindled on carrier I56; the pinion I5I meshwith a set of planets I58 spindled on carrier Annulus gear I51 meshing with I54 is rigidly joined to rotate with carrier I59. Annulus gear I6I is attached to drum I62, serving as the reaction member of the second unit. Drum I62 likewise supports the set-of clutch plates I60 and a prmser plate I63, similarly to drum I8 of the first unit, and has integral cylinders I65 accommodating pistons I66, engageable by fluid pressure supplied from gland 35 through passage 33'. Clutch release springs are shown at I61. The fluid pressure for clutch pistons 32 is fed through pipe I through passages 34 and 33 to cylinders 3|.

' Carrier I56 of planets I54 drives the loadshaft or tailshaft 300. Brake I10, like brake 30 of the first unit, is operable by fluid pressure and spring storage as shown in Figure 4. Drum I62 is supported on shaft 300 by bearing 30I.

An extension of drum I62 mounts aflixed input sun gear "I of the reverse driving or third unit. Carrier I12 fixed to tailshaft 300 has spindled planets I13 meshing with sun gear HI and with annulus I15 having drum I16.

For reduced forward drive in the second unit, brake I10 is locked; for direct, clutch I55-I60 is engaged and brake I10 released.

later, the ratio drive com- First unit Second unit Low Brake 30 Brake 170. Clutch 15-20. Brake 170.

Brake 30 Clutch 155-160. Clutch 15-20. Clutch 155-160.

To obtain reverse drive, the first unit brake 30 is energised, the second unit brake I10 and clutch I55I60 are both released, and brake I80 of the third unit is locked to annulus drum I16.

It will be seen that if a slow forward speed be imparted to shaft I0, with shaft 300 and carrier I56 stopped, sun gear I52 will rotate annulus I51 backwards or reversely to the direction of input torque. This causes a reverse speed component on carrier I59.

At the same time, sun gear I5I will be imparting a backward component to annulus I6I if carrier I59 were stopped; but, carrier I59 is not stopped, but has a reverse component, so that annulus I6I is being driven reversely at increased or accelerated speed.

Since brake I80 has locked annulus gear I15, the reversecomponent applied to sun gear "I by annulus I6I through drum I62 is demultipl'ied in the third unit, so that carrier I12 applies a reduced ratio reverse component to shaft 800.

This method of obtaining reverse drive at low ratio is believed novel, and provides a new result in the derivation of negative components from positive directional torque, in compounded gearing.

For the purposes of the present application, it is not deemed necessary to show other than a conventional form of control for the operation of units II and III.

The disclosure of Figure 7 Is of the actuation system for unit II, for brake I10 surrounding drum I62. Brake end I10a at the left is pivoted to thrust rod I8I fitting notch I82 of bellcrank lever I83 pivoted to the casing 2 at I84. Brake end I86 is the fixed anchor point, and is held by adjustable stud I81.

Servo cylinder I90 controls the application brake I 10. Servo line I95 likewise feeds clutch cylinders I65 of Figure 1. Piston I88 attached to brake applying rod I89 is loaded by springs I9Ia, Hill), and I9Ic for normal actuation. At extreme rightward motion, abutment I92 sliding on rod I89 may deliver the force of spring I9Ic to pins I99 fixed to the second piston I94 and protruding through holes in piston I88; thence to piston I94 sliding on rod I89 in sub-cylinder I96.

Pipe'l91 connects cylinder I90 to port 226 of valve W of Figure 8. When pressure is admitted to cylinder I90, piston I88 shifts to the right, compressing springs I9Ia, I9"), and I9Ic, and relieving brake I10 from engagement with drum Pipe I98 connects to port 238 of valve 235 of Figure 8 so that compensating pressure from line I00 and valve 235 may act on piston I94 to vary the pressure build-up in cylinder I90 reacting through lines H and I95 11 clutch 55-450 when the control valving of Figure 8 is moved to establish drive therein.

Figure 8 is a schematic representation of the entire control system. Governor G and pump P are shown connected to the valving of easin wherein the starting control of unit I is derived; to the individual valves V V V, which select ratio drive for units I, II, and III, by directing the fluid. pressure of pump P to the servo cylinders of each for clutching and braking; and to the compensator valve C moved by the operator's engine speed control pedal or accelerator, to be described later.

The selector valves are all of the three-port balanced type, V controlling the pump pressure from pump line II5 to cylinder 60 through pipe IN and to clutch cylinders 3| of the first unit. Valve V controls in the same manner thepump pressure from line 5' to cylinder I90 and to clutch cylinders I65 of the second unit. Valve V similarly controls the pressure delivered to cylinder 2I0 arranged to load and actuate band I80 of the third unit through thrust rod 2I2 and piston 2I3. All three valves are shown in exhaust or draining position wherein brake I10 of the second unit is applied by springs I9I, and brake 30 of the first unit may be manipulated by valve I 20 of Figure 4.

The servo cylinder 2I0 for the brake I80 of unit III is shown in section in Figure 8, the lever 2 being arranged to actuate the band I80 in the same manner as lever I83 of Figure '7 actuates band I10. Rod 2I2 moving in cylinder 2I0 may load band I80 through lever 2 when fluid pressure is admitted to cylinder 2I0 behind pis-. ton 2I3 through pipe M5; and the pressure being relieved, springs 2I4a and 2I4b may release the band I10, since piston 2I3 sliding on rod 2I2 and striking abutment 2 I6 thereof, may transfer energy between rod 2I2, springs 2I4a and 2I4b and the fluid pressure of cylinder 2I0. Weak spring 2I8 provides initial response .to pressure build-up in the cylinder 2I0 before piston 2I3 strikes abutment 2I6 so that an initial brake loading pressure of predetermined value is available. It also prevents full retraction of piston 2I3 under the loading of springs 2I4a and 2I4b, so that piston H3 is under constant stress at all times. The-lost motion between abutment 2I6 and piston 2I3 assures that when the fluid pressure is released from cylinder 2I0, there is no direct interaction between the band I80 and the piston 2I3, so that residual torque drag is allowed to dissipate and the band I80 free itself without interference from the fluid pressure system.

Valve V has pressure port 2 2I connected to pump line II5 servo port 222 connected to pipe IOI for delivering pressure to piston 63 in cylinder 60 of Figure 2. Port 223 connects to exhaust.

The pressure port 225 of valve W connects to pump line II5, servo port 226 to line I95I91, and cylinder I90 of unit II, and port 221 is open to exhaust.

Valve V has pressure port 228 Joined to pump line H5, servo port 229 connected to line 2I5 of cylinder 2I0, and exhaust port 280. Valve V has pressure port 2, servo port 242 leading to line I95 and clutch cylinders I65 and exhaustport 243; biasing spring 244, and pressure shift port 245 connected to line 2I5.

The valves may be conveniently located in a common casing such as 220 (Fig. 10), along with valve 0, as space considerations permit.

The three ratio selector valves are, for the purposes of this specification, manually operated for the four forward and reverse gear ratios.

The valve V is for the purpose of permitting valve W to hold off brake I10 when reverse drive is desired, but to prevent clutch I55--I60 from being actuated. Valve V when shifted to admit pump pressure to cylinder 2I0 for applying reverse band I80 to drum I16, also delivers pressure to shift valve V Port 2 of valve V is connected to line I95-1, port 242 is connected to line I95, and clutch cylinders I65 of the second unit, and port 243 is open to exhaust. Spring 244 normally compels the valve to remain in the down position, so as to connect ports 24I-242, lines I95-I95, and permit valve V to operate simultaneously on brake piston I88 and clutch pistons I66. This is for forward drive. When valve V is shifted to reverse position, fluid pressure from pump line I I5 passes through line 2I5 to cylinder 2I0, loading piston 2I3 and brake I80 against the action of springs 2M; and also enters port 245 of casing 220 lifting valve V against spring 244 and cutting off clutch loading pressure from line I95 to cylinders I65, connecting the latter to exhaust.

For reverse shift, therefore, brake 30 of unit I is gradually applied, and brake I80 of unit III. Neither brake I nor clutch I55I60 of unit II is actuated.

The automatic cutting out of clutch I55-I60 while brake I10 is held off, when the reverse shift valve V is active, is believed a novel feature. As soon as valve V is returned to a draining position spring 244 reconnects valve V to control over clutch I55--I60 of the second unit, so that there is a constantly eifective fluid pressure interlock preventing wrong motion of the controls.

It will be seen that the shift pattern of the valving matches the ratio table given preceding in the following manner: (This symbol 0 indicates corresponding valve admission of servo pressure.)

providing actuation of the clutches and brakes designated:

Reverse Neutral Low 2nd 3rd High Applied Applied Applied Applied Applied In the above table, it will be noted that in reverse, the valve V is admitting pressure to lines I95-I91 to hold ofi brake I10, but valve V is furnishing pressure to maintain V in a position to drain cylinders I65, and compel disengagement of clutch I55-I 60 of unit II also.

In neutral it is preferable that the path of torque be broken in unit I, therefore the fluid control is so arranged that while springs 18 may be loading brake 30, the lack of fluid pressure in cylinder 16 has eliminated brake end 36 as an anchor point. When the engine is started, the load of the vehicle already applied to shaft I0 and carrier 1 may serve as a reaction means, and the spinning of annulus gear 5 causes the sun gear U and connected drum III to spin reversely to the direction of rotation of the engine shaft 3.

If the operator were to race the engine by'depressing the accelerator pedal, governor G might deliver a pressure of a magnitude that valve I20 of Figure 4 would be moved to initiate drive in unit I,

It is believed a proper measure for increased safety to establish a positive no-drive neutral elsewhere in the assembly, so that an inexperienced operator may not ever have a drifting of the vehicle under power, therefore the valving for controlling unit II is arranged to give a positive neutral condition when the manual control is put in neutral.

It will be noted that in Figure 8 valve V has an extended stem protruding through the housing with a lip or flange for engaging with the external control lever such as shown in Figure 9, so that it can be moved manually against the action of weak spring 244. This is so that valve V can admit pressure to line I95-I91 and cylinder I90 to hold out brake I10 of unit II, and at the same time, fluid pressure in line I95 can be cut off, and clutch cylinders I 65 drained, when no-drive in unit II is desired, as in neutral. The manual means to shift valve V is described later in conjunction with Figure 9.

The coordinating of the various positions of the selector valves may be done in many ways, according to the desires of the designer and the requirements of the installation. One method of establishing the foregoing shaft pattern is shown in Figure 9.

Cam plate 250 rotating on shaft I is moved by rod 252 pivoted to arm 253, the rod 252 being moved by lever 202 and shaft 20I from hand lever 200 mounted to reciprocate over sector plate 2I0, on the steering column of the vehicle, as shown in Figure 10.

The showing is conventional, there being two slots out in cam plate 250, each guiding a pin of a lever arranged to shift one of the valves. On the opposite side from the slots, the cam plate 250 is cut to a contour 255 such that two additional pins of two additional levers may rock the levers according to a predetermined pattern corresponding to the ratio shift requirements of the preceding tables.

Referring to Figure 8, the four valves of that diagram are shown in Figure 9 protruding from the casing 220, each to be worked by a lever.

The valves are moved by individual levers mounted to rock on the common rockshaft 249 of Figure 9.

It should be noted that when valves V V and V are in their lower positions, the pump line H5 is connected through to the actuators for the units controlled, and when in their up positions, the pump line H5 is closed off and the unit concerned is not actuated by the fluid pressure, but connected to exhaust. Valve V in the up position connects clutch cylinders I65 to exhaust, and in down position connects pressure line I95 to the clutch cylinders I65, when pressure exists in line I95 by virtue of a down setting of valve V These respective up and down positions are identical in Figure 9 so that valve V connected to lever 260 is in the up" position when pin 36I occupies the longer radial distance from the center of shaft I in slot 254 for the ratio positions of reverse, neutral, low and 3rd, and is in the down position when pin 26| occupies the shorter radial distance from the center of shaft 25!, for the positions of 2nd and High.

Valve V attached to lever 262, is down when pin 263 occupies the longer radial position in slot 256 in reverse, neutral, 3rd and High; and is up when it is in the shorter radial positions for low and second.

Valve V moved by lever 26! and pin 266 is down when pin 266 is at the longer radial positions of reverse and neutral, and up" for all other settings. Sprin 26'! loads lever 266 so that pin 266 will follow cam 255 at all times.

Valve V moved by lever 265 and pin 268 is spring loaded by 260for following the cam 255 and is up for reverse but down for all other settings of camplate 250 and lever 253 moved by rod 252 and the linkage of Figure 10 to the handlever 200.

These up and down positions correspond to the tables of ratio shift given preceding so that the operator of the vehicle may establish a manual shift for any of the required conditions by merely moving handlever 200 over sector plate M0 to any ratio designation, and the servo actuation system comprising the elements of the showing of Figures 1 and 8 will immediately establish the selected ratio.

In Figure 10 the view of the steering column and steering wheel of a vehicle shows the operators handlever 200 pivoted to reciprocate over a sector plate 2|0 attached to the column, and marked R, N, n2," 3, and n4 for ratio shift positions to be determined by the servo devices described. Button 200a operates a common poppet mechanism (not shown) for the correct positioning of the lever with respect to the plate, as well as for the controlled selector mechanism.

Lever 200 rotates rod 20l and attached lever 202 pivoted to rod 252, which reciprocates arm 253, shaft 25l and selector cam plate 250 shown in detail in Figure 9. The positions shown in Figures 9 and 10 are for the selections of reverse speed ratio.

The operator's hand control lever 200 is mounted on appropriate brackets so as to rotate shaft 20| with reciprocation of the handlever. The lower end of the shaft 201 is rigidly attached to a curved arm 202 pivoted to rod 252. The rod transmits the efiort of the handlever 200 to shift the valving, which selects actuation of the ratio changing elements herelnbefore described.

Lever 253 is attached to shaft 25! mounted in the housing 220 and shaft 25! is attached to cam plate 250 in which is cut slots 254 and 256 accommodating pins 26l and 263 of levers 260 and 262 pivoted to the housing at shaft 249. The opposite ends of the levers are forked so as to engage the upper ends of valves V and W for moving the valves for distributing the fluid pressure from the pump line H5 to the various ports controlling the fluid pressure servo system of Figures 4, 8, and 9.

This is a purely manual control for establishing speed ratio in the various combinations noted preceding. It is within the purview of my invention to control the movement of the valving by automatic means responsive to changes in speed applied to a. second governor, similar to governor G shown in Figure 4, deriving speed from either of shafts 3 or 300 or any other rotat ing part of the mechanism.

The lower part of Figure 10 shows engine accelerator pedal 205 arranged to exert a downward thrust on rod 200 rocking lever 20'! of cross-shaft 206 when depressed. Lever 204 of shaft 206 is connected to the engine throttle rod 203, and lever 209 is pivoted to rod 231 for operating the compensator valve C, having the number 235 in Figure 8., The customary accelerator pedal retracting spring is shown at 206a.

To this point, no detailed explanation has been given of the auxiliary control of valve 235, and the operation of the lower part of the structure of Figure 10.

Figures 8 and 10 show rod 23| which is moved to the left with increase of engine throttle opening, as the engine accelerator pedal 205 is depressed. The rod 23l is pivoted to lever 232 mounted on shaft 233 recessed in the casing,

and as the pedal is depressed, lever 232 attached to shaft 233 rotates counterclockwise.

The valve 235 sliding in bore 236 of valve casing 220' is integral with stem 235a, which is moved by spring 231 recessed in plunger 240, which is engaged by the lever 232' noted preceding.

The plunger 240 is drilled out to accommodate the enlarged end of stem 235a so that after given range of motion of lever 232 the stem may be solidly engaged by the lever, instead of the force applied to valve 235 being delivered through spring 231.

The porting in order from the top of the drawing of Figure 8 consists of port 234 connected to pipe H5, port 238 connected to pipe I00, port 239 joined by drilled passage to exhaust excess pressure to the sump, and counterport 238a joined to port 238 by a side passage.

The valve 235 positioned by the movement of accelerator pedal 205 and by pressure in line H5 exerted through port 234 on the upper face of the valve 235, is a metering device for regulating the degree of pressure in line [00, which is connected to cylinder 66 of the control arrangement of Figure 2.

It will be understood that changes in pressure delivered by pipe IM to cylinder 6! are exerted on the normal brake locking action of springs 18 acting through piston 61 on brake pressure rod 50.

When rod 58 is moved to the left under pressure from pipe IN, the first action is for spring 19 to be compressed before the abutting end of the sleeve of piston 63 can strike abutment 65', which is locked solidly to shaft 58. When the abutting action occurs, the value of whatever fluid pressure is maintained in cylinder 66 through pipe I00 is exerted against springs 18, so that the building up of pressure in pipe l0| for a given linear movement of rod 58 to brake release position is proportional to the pressure in cylinder 66'. If the pressure in 66 acting on piston 65 is high, the brake 30 will become disengaged at a low line pressure which is also acting on clutch |520. Since the movement of the accelerator pedal is regulating the compensating pressure resistance in cylinder 66, the operator may, by sharply depressing the accelerator pedal, quickly decrease the compensation pressure so that the full effective pressure on clutch l5--20 can provide maximum torque capacity. At relaxed throttle, the net servo line pressure being exerted on clutch l5-20 through pistons 32 in cylinders 3| may be as low as onethe compensator pressure press springs I 9Ia and I9Ib as Well as fourth of maximum pressure, giving the driver a range of graduated pressures for meeting any traffic conditions for heavy or light torque demand.

When pedal 205 is depressed, the compensator line I00 rapidly loses line pressure, and'when the pedal is relaxed'to decelerate the vehicle engine, in I00 and connected cylinders I96 and 66 buildsup.

The purpose of the driver-will control over the compensator pressure is to aiford graduated regulation of the engaging pressure on clutches I520 and I55-I60 when a ratio shift selection of valves V or V is made in which pump line pressure from pipe I I5 is admitted to either of lines IM As noted preceding, the admission of pressure to line IOI sets up a force tending to store energy in brake springs 18 of Figure 2, the net pump-tosprings resistance plus that of clutch pistons 32 and spring 22 of Figure 1 determining the effective loading pressure on the clutch plates. Now when the resistance of springs 18 is already partially canceled, the loading pressure being sustained on the clutch I5-20at the interval of release of brake 30 is going to be less than that when the full resistance of springs 18 is felt by the pump pressure. At full throttle the clutch torque capacity is therefore at maximum, while at relaxed throttle it is at some predetermined smaller value.

In the same manner, the variationsof pressure in compensator line I00 are applied to piston I94 of cylinder I96 of Figure 7 through line I98. The piston I94 may act directly on piston I88 to comthrough pins I99 take out springs I9Ic through their striking abutment I92.

' The feature of compensator action which enables the driver to adjust over a continuously variable scale of pressures, the net driving torque of the friction elements being established for driving is believed novel in combination with the other features of this disclosure, wherein selected shift between any of the ratios may occur under all driving conditions.

Attention is directed to the arrangement of gears between shafts I0 and 800, whereby a. reduction reverse gear drive is obtained by multiple compounding of the gearing in the second and 7 third units.

Assuming a forward driving component applied to shaft I0 and pinions I 5I ,and I52, if the output shaft 300 be stopped, the cage I55 will be stopped, and forward rotation of pinion I52 will impart a reverse component to annulus I51.

Now if it be assumed that cage I59 be stopped, pinion I5I, would impart a reverse component to annulus I6I through planet I58. But cage I59 is not stopped but already has an imparted reverse component from the interaction of pinion I52. annulus I51 and'pinion I54, therefore the resultant of these motions on annulus I8I and drum I62 is an accelerated reverse motion.

Sun gear "I of the reverse unit therefore re- 1 ceives a reverse component, demultiplied when brake I is applied, to give cage I12 connected to output shaft 300, for reverse drive inreduction. Brake I10 is, of course, not applied when driving in reverse.

This arrangementis believed novel in the art, and affords useful demultiplication of drive in reverse obtainable otherwise only through more complicated, expensive means, whereas the arrangement herewith is neat and compact.

In Figure 11 a fragmentary view of the servo cylinder 2I0 of Figure 8 is given, the plunger 2I2 projecting from the-cylinder to coact with lever 2I I and thrust rod 215 which-engages band I80 in the same way that thrust rod I8I engages band I10 of the second unit.

As will be understood from the preceding'de scription of the gear relationships of the second and reverse units, if both of brakes I10 and I80 be simultaneously applied, a locking couple will be established between drum I62, sun gear "I and annulus I15, preventing rotation of cage I12 and shaft 300.

In referring back to the control relationships described preceding in connection with Figures 3, 4 and 8, one will note that if the engine should stall with the car drifting backward downhill, there would not be any engine braking, since the stalled engine would prevent the pump of Figure 3 from operating, and without pump pressure in main II5 of Figure 8, the band I10 of the second unit would be applied by springs I9I of Figure '1, however, the reversal of the direction of torque by backward drifting of the vehicle would tend to de-energize band 30 of the primary unit because of the wrapping direction withrespect to arrow Y of Figure 2, hence the torque'forces militate against engine braking under these circumstances.

Therefore the structure of Figure 11 includes rod 2" pivoted to lever 210 rocking on spindle 212,

and engaging stop 214 of lever 2 at 213.

The rod 2" is normally in a left-hand position such that portion 213 of lever fere with the normal control of plunger 2 I2 over the action of brake I80 of the reverse unit.

When, however, a rightward thrust is applied to rod 21I, lever 210 rocks clockwise, and through stop 214, applies a force through notch 2160f lever 2 on thrust rod 215, pivotedto brake I80 of Figure 1. The application of brake I80immediately sets up a locking couple through the gears as above described, so that the car driver is equipped with supplementary braking means other than the vehicle brakes to take the place of engine braking, and yield a positive safety means against the car getting out of'control in mountainous country.

Rod 2'" is attached to the ordinary emergency brake lever of the'present day vehicle, to remain in the left-hand position aforesaid when the emergency brake lever is in theoif position; and to be thrust to the right when the emergency brake is applied, to apply the brake I80 of the reverse unit.

It is not deemed necessary to show the connection of rod 2" to a lever or pedal operative upon the vehicle brakes, such construction being common and well understood. This feature is believed novel, and useful, and is not known to the present applicant as existing in the prior art. It relates to all forms of transmission devices in which there is a loss of engine braking when the engine is stalled, since it provides supplementary braking through a locking couple within the gearbox, controlled by the vehicle braking controls. It is within the purview of the invention to connect rod 2" to an entirely supplementary .lever, manually or otherwise operated independently, rather than through the control for the vehicle braking system, such as a pedal in the position of the customary main clutch pedal, now dispensed with in the present invention. Such a pedal may be labeled noof Figure 7 210 cannot interto use it for driving in hilly load said member for back" or "anti-reverse", the driver being taught country.

Although a specific embodiment of the invention has been illustrated and described, it is to be expressly understood that it is not limited thereto, but that various changes may be made in detail, design and arrangement without departing from the spirit and scope of the invention.

Reference is therefore directed to the appended claims for proper definition of the scope of the invention, wherein I claim:

1. In power control devices for motor vehicles, a power shaft and a load shaft, a change speed gearing embodying a speed ratio determining member, a governor driven in accordance with the speed of one of said shafts, fluid pressure means controlled by said governor effective to actuate said member, including valving for varying the value of and directing the fluid pressure, and control means responsive to driver-determined action operative to establish graduated actuation of said member by said fluid pressure means, whereby smooth initial transfer of torque between the shafts is obtained.

2. In power controls, a motor vehicle having power and load shafts connected to a primary gear unit including manual controls for said unit, and with a secondary gear unit including automatically operated speed responsive controls therefor; a control system arranged to establish initial actuation of speed ratio in said last named unit according to speed response of said automatically operating controls, and connecting means operative between said system and said manual controls whereby initial selection of drive in said first named unit renders said system and said automatically operating controls effective to establish said actuation in said last named unit for starting the vehicle from rest.

3. In power controls, a motor vehicle embodying a change speed gear unit, a friction member normally arranged to establish speed ratio change in said unit, additional friction members arranged to establish a different speed ratio from that of said first named member, actuating means for said members, fluid servo means controlling said actuation, ratio changing controls effective upon said fluid servo means whereby selective shifts of speed ratio are obtained by selective actuation of said members, and separately controlled means separately acting upon said fluid servo means effective upon one of said members to controllably actuate initial drive gradually in starting the vehicle from rest.

4. In automatic power control devices, a motor vehicle having power and load shafts, a variable speed gearing coupling said shafts including a speed ratio determining brake band; a speed responsive governor, manually controllable means normally active to control the action of said band for establishing speed ratio in said gearing, and separate control means connected to said governor effective to set aside the first named means, and control the action of said band for initiating drive between said shafts.

5. In power control mechanism, a variable speed gear having a rotatable reaction element, a friction member subject to the directional torque of said element, action and reaction sustaining devices for said member, fluid pressure actuation means for said devices arranged to stopping the rotation of said element, and selectively operated control selective operation of said means effective upon said actuation means for devices for action or reaction according to the direction of torque of said element.

6. In power control devices, a change speed gearing having a rotatable reaction element, a friction member mounted for limited rotational motion with respect to the rotation of said element a pressure exerting device normally biased to load said member and displaced with respect to the idling direction of rotation thereof when said load shaft is stopped that said member is energised for engagement by force applied to it from the rotation of said element, a second pressure exerting device arranged to load said member and displaced such that it may act oppositely to the action of said first named device, and controlmeans effective to select actuation of said devices whereby initial actuation of said member on said element is obtained without rotational energisation of said member by said element.

'1. In a transmission braking device, for establishing speed ratio, an input shaft, an output shaft, gearing intermediate said shafts, a braking member biased for disengagement from a reaction drum, a movable anchorage for said member displaced with respect to the normal reaction rotation of said drum to yield a self-wrapping action of said member when energised, a second movable, anchorage for said member displaced with respect to the normal reaction rotation of said drum to yield non-wrapping action when energised, and control means adapted for selective energisation of said anchorages whereby non-wrapping or selfwrapping action of said member is selectively obtained.

8. In transmission ratio control devices, an input shaft, an output shaft, and change speed gearing coupling said shafts including reaction supporting means for establishing drive through said gearing between said shafts, a braking member for said means, and movable anchorages for said member oppositely disposed with respect to rotation of said means, actuation mechanism for said member effective upon said anchorages, selector controls arranged to establish selective actuation by said mechanism on said anchorages and upon said member, and a supplementary manual control to said selector controls operative to regulate the degree of actuation by said mechanism on one of said anchorages whereby graduation of the force of said member upon said reaction supporting means is in accordance with the movement of said supplementary manual control.

9. In power controls for motor vehicles, a transmission gearing system comprising a power shaft, a servo pump driven by said shaft, three transmission units arranged in series and a load shaft, controls for the output pressure of said pump adapted to shift the speed ratios of said units for forward, neutral, reverse, and a plurality of forward speed ratios, at least one of said units including friction driving means operated by fluid pressure, and a supplementary manual control for the output pressure of said pump arranged to apply a graduated fluid pressure to said friction driving means for initiating the drive between said shafts when said load shaft is not rotating.

10. In power controls for variable speed gearing, a power shaft and a load shaft, a servo pump driven by one of said shafts, a plurality of trans mission gear units arranged sequentially between the said shafts, a plurality of friction elements for establishing various speed ratios of drive by said units between said shafts, a plurality of fluid servo devices for actuating said elements by fluid pressure derived from said pump, valving arranged to distribute and release fluid pressure derived from said pump to and from said servo devices, and supplementary servo control valving for making the pump pressure eflective upon at shaft is not rotating.

11. In power control devices, a power shaft and a load shaft, transmission mechanism adapted to transmit forward and reverse drive between said afts,

12. In transmission controls for motor vehicles, a power shaft, a load shaft, gearing coupling said 15. In power control devices for motor vehicles, a powershaft speed gear unit therebetween having a reaction member for establishing drive through said gearing, a friction element adapted to prevent rotation of said member and to connect the shafts for drive at initial driving speed ratio for starting the vehicle from ing said power shaft.

17. In power handling mechanisms, a motor vehicle having a power shaft and a load, shaft,

actuation of said devices, governor coacting with said means, and a supplementary manual control coacting with said governor and said means operative to establish initial drive by one of said elements for either forward or reverse drive of the vehicle.

to graduate the driving engagement for, the one said ratio controlled by said governor.

19. In motor vehicles, a power shaft and a load shaft, a step ratio gearing between the shafts, a plurality of friction tablish a series of step-ratio drives in said gearing, a plurality of actuators for said elements including one actuator effective to initiate geared drive only between said shafts through one of said elements, fluid pressure responsive means ,operative upon said actuators, a pump supplying pressure to all said actuators, including that actuator utilized to. initiate drive between. the

shafts, and speed operated valwng adapted to pressure from said pump to said fluid pressure responsive means.

20. In a combination such as described in claim 19, the subcombination of a further manual control member coacting with said valving for varying the regulation of the flow of said pressure to said fluid pressure responsive means.

1 L. CARNEGIE.

elements adapted to es- 

